Linear pump with exhaust pulsation attenuation

ABSTRACT

A linear pump having an axially aligned cylinder and piston arrangement driven by an electromagnet motor has an exhaust chamber defining a cavity covered by a diaphragm. The diaphragm is movable toward the cavity in response to pressure fluctuations in the exhaust chamber to attenuate pulsations in an air stream exiting the exhaust chamber. A diaphragm is mounted over the recessed cavity by a support ring having an open center allowing air to act against the diaphragm.

CROSS-REFERENCE TO RELATED APPLICATIONS

Not applicable.

STATEMENT OF FEDERALLY SPONSORED RESEARCH OR DEVELOPMENT

Not applicable.

BACKGROUND OF THE INVENTION

The present invention relates to pumps and in particular to compact linear piston pumps.

Pumps for certain duties, such as oxygen concentration and sewage aeration, generally need to be compact and operate discreetly. It is thus important to properly muffle the working air as well as reduce vibration during operation of the pump without relying on a large, thick-walled housing to attenuate the sound and vibration. Discreet operation of the pump can be obtained by insulating the housing, however, this adds bulk and can cause cooling problems. Mufflers can be added at the output, however, this adds hardware and cost.

Such compact linear pumps/compressors are often single cylinder devices with a small piston that reciprocates rapidly within a small cylinder to pressurize the air. The rapid movement of the single piston generates considerable vibration. These vibrations are often transferred directly to the pump housing, via a direct rigid mounting connection.

To facilitate reciprocation of the piston with less vibration, it is known to suspend the drive member, such as the armature of an electromagnetic motor, by springs or like flexible members. Stacks of thin metal leaf springs are well-suited for this. However, when multiple springs are used, it can be difficult to achieve the prescribed spring rate to which the piston drive components of the pump have been tuned. Changes in the angular (about the piston axis) and/or axial (along the piston axis) orientation of the springs relative to one another can effect the spring rate. To ensure that the pump operates efficiently, it is thus important to achieve the intended spring rate and thus ensure the consistent orientation of the suspension springs, which can make pump assembly difficult.

Another problem is that the intake and exhaust valves of the valve head must open and close rapidly for each stroke of the piston. Typically, thin metal flapper valves are used for this purpose because of their ability to seat and unseat very rapidly. Since the exhaust port opens under the force of the compressed air, a valve stop is used to support the valve and prevent it from being hyper-extended beyond its elastic range. The rapid contact between the intake valve and the valve head or the exhaust valve and the valve stop can generate tapping or clicking sounds. Another problem is that the rapid opening and closing of the intake valve can cause pressure fluctuations or pulsations in the air flow upstream from the valve head. These air pulsations can generate a low-frequency, rumbling noise.

Another problem confronting the design of compact linear piston pumps is eliminating pulsations in the output air stream. Pulsations in the air downstream from the outlet has been found to alter the resonant frequency of pump when different lengths and/or diameters of output lines are attached to the pump. Changing the operational frequency of the pump causes inefficiencies that can ultimately render the pump unusable for particular applications. It can also exacerbate noise and vibration issues.

Yet another persistent problem in compact linear pump design is cooling. To decrease noise, or perhaps to make immersible or suitable for outdoor use, the working components of these pumps are often enclosed in a pump housing. With operation of the pump, friction and the current in the electromagnet coil generate heat. As is well understood, heat adversely affects the pump efficiency and life. Many times the need to keep the pump operating efficiently requires the housing to be vented or to have other measures taken which destroy, or at least significantly reduce, the noise retarding features of the housing or other components.

Accordingly, an improved linear pump is needed that addresses the aforementioned problems.

SUMMARY OF THE INVENTION

The present invention is an improved linear pump having a pulsation attenuating mechanism in its exhaust chamber to provide a smooth air flow out of the pump.

In particular, the linear pump includes a cylinder and piston arrangement and an electromagnet with a stator containing a wire coil driving an armature to reciprocate the piston within the cylinder along a piston axis. The cylinder is in communication with the exhaust chamber through an exhaust port of a valve head. The exhaust chamber includes an attenuator movable in response to pressure fluctuations in the exhaust chamber to attenuate pulsations in an air stream exiting the exhaust chamber.

In preferred forms, the attenuator is a diaphragm, preferably in the form of a rubber disc mounted over a circular cavity by a plastic diaphragm support having an open center, which allows air in the exhaust chamber to act against the diaphragm directly.

The linear pump of the present invention thus outputs smoother air flow by decreasing the amplitude of air pulsations in the exhaust chamber that arise principally by virtue of the single cylinder arrangement. By reducing pulsations in the air downstream from the outlet the pump becomes less sensitive to attachments downstream from the pump outlet. Thus, the resonant frequency of the pump can be maintained within an acceptable range for a variety of lengths and diameters of the output hose or tubing. The linear pump of the present invention can thus operable more efficiently for a greater number of applications.

These and other advantages of the invention will be apparent from the detailed description and drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a perspective view of a linear pump according to the present invention;

FIG. 2 is a perspective of the pump of FIG. 1 shown with a housing shroud removed;

FIG. 3 is an exploded assembly view of the pump;

FIG. 4 is an exploded assembly view of a drive components of the pump;

FIG. 5 is a sectional view taken along line 5-5 of FIG. 6;

FIG. 6 is a top view showing the pump with an inlet filter and top cover removed;

FIG. 7 is a partial sectional view taken along line 7-7 of FIG. 6 albeit with the inlet filter and top cover in place;

FIG. 8 is a top view of the pump with the housing shroud removed and the drive components shown in phantom;

FIG. 9 is a top view of a pump base showing the intake and exhaust chambers thereof;

FIG. 10 is a sectional view take along line 10-10 of FIG. 9;

FIG. 11 is a partial sectional view taken along line 11-11 of FIG. 9;

FIG. 12 is a sectional view taken along line 12-12 of FIG. 1;

FIG. 13 is an enlarged partial sectional view taken along arc 13-13 of FIG. 12;

FIG. 14 is a sectional view taken along line 14-14 of FIG. 2 albeit shown with the housing shroud in place;

FIG. 15 is a partial sectional view taken along line 15-15 of FIG. 2 albeit with the outer housing in place;

FIG. 16 is a view of a retainer ring shown in isolation as viewed from line 16-16 of FIG. 12; and

FIG. 17 is a view of a leaf spring shown in isolation as viewed from line 17-17 of FIG. 12.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT

The present invention provides an axial or linear piston pump. The term pump used herein includes a device for providing either positive or negative pressure, and thus either acting as a vacuum pump or a compressor. The pump has a compact form factor, with a preferred operating range of 2-30 psi, depending upon the application (however, the pump could be designed to operate at other pressures) with low external vibration and noise and less sensitivity to pump attachments (lines, hoses, tubing, etc.) downstream from the outlet.

Referring to FIGS. 1-3, the pump, generally referred to in the drawings by reference number 20, has a compact housing 22 including a base 24, a shroud 26 and a top cover 28. The shroud 26 is bolted to the base 24 (via bolts 23), with a gasket 30 therebetween (see FIG. 15), and the cover 28 is bolted (via bolt 69) to the shroud 26. The base has four feet 32 to support the pump 20 and provides an opening for connecting a fitting 34 (see FIG. 11). The shroud 26 has an opening 35 for an electrical socket (or power cord). The cover 28 is spaced off of the shroud 26 slightly to define an air intake passage 36 along the periphery of the cover 28, as will be described in greater detail below. The top of the shroud 26 and the cover 28 are recessed on opposite sides to provide a hand gripping area 38.

Referring to FIGS. 3, 5-7 and 12, the pump 20 breathes through an air intake assembly generally designated 40. The air intake 40 is configured to reduce the low, rumbling noise associated with pulsations in the intake air caused by rapid movement of the intake valve. In particular, the air intake 40 includes an inner cavity 42 defined in part by a recess in the top of the shroud 26, a seal partition 44, a filter tray 46, a filter 48 and the cover 28, which provides the upper boundary for an outer cavity 50.

As mentioned, the recessed top of the shroud 26 defines the inner cavity 42, in the floor of which are two spaced apart orifices 52 near one end (the right end in the drawings). The inner cavity 42 is bounded at the top by the partition 44 which seals against a peripheral wall 54 of the inner cavity 42. The partition 44 has a pair of spaced apart orifices 56 located at a (left) end of the inner cavity 42 opposite the orifices 52. To facilitate assembly the partition 44 includes another set of orifices 56′ at the opposite end, however, these are not used when the partition 42 is oriented as shown in the drawings. In any event, intake air flows through only the one set orifices 56 (or 56′), which is located opposite the orifices 52 in the floor of the inner cavity 42. Resting on the partition 44 is the filter tray 46 which holds the filter 48 in the outer cavity 50. The filter tray 46 has a bottom wall with a pair of openings 58 which align with orifices 56 in the partition 44. A small alignment feature 59 extends down from the underside of the filter tray 46 and fits into an opening 61 in the shroud 26 to ensure that the filter tray 46 is assembled in the proper orientation. The filter 48 is held spaced off of the bottom of the filter tray 46 by a number of small spaced apart risers 60, and is retained by a short peripheral wall 62 ringing the filter tray 46. The wall 62 has a cut-out 64 at one end (opposite the openings 58) allowing intake air to flow laterally into the filter 48. The cover 28 fits onto the shroud 26 over the filter 48 to define the outer cavity 50. Ribs 66 on the inside of the cover 28 contact the wall 62 of the filter tray 46 to keep the bottom edge of the peripheral wall 68 of the cover 28 spaced slightly from the shroud 26, and also to funnel the intake air into the outer cavity 50 through the cut-out 64. The space between the top cover 28 and the shroud 28 defines the intake air passage 36, which extends around the periphery of the cover 28. Aligned center openings in the cover 28, filter 48, filter tray 46 and partition 44 allow a bolt 69 to screw into a threaded opening 70 in the shroud 26 to secure the assembly.

Referring now to FIGS. 5, 7, 8 and 12, the shroud 26 also integrally defines a pair of vent tubes 72 which extend down form the orifices 52 in the floor of the inner cavity 42. The vent tubes 72 inject the intake air passing from the intake 40 to the inside of the shroud 26 past the drive assembly, described in detail below, through three openings 80 in a base plate 82 (and elongated opening 84 in the gasket 30), which is bolted to the base 24, and into an intake chamber 86 in the base 24 of the housing, as shown in FIG. 5. Note that the base plate 82 is bolted to the base 24 by bolts 87 (see FIGS. 2 and 3).

As shown in FIG. 9, the intake chamber 86 is defined in the base 24 with an exhaust chamber 88 in a yin-yang-like configuration, each with a wide portion and a narrow portion in opposite orientation. The intake air is drawn from the intake chamber 86 up through an intake opening 90 in the base plate 82 (and an associate opening the gasket 30) and an intake tube 92 connected to the intake side of the drive assembly, as shown in FIG. 10. Exhaust tube 94 runs between an exhaust opening 96 in the base plate 82 (and an associated opening in the gasket 30) to the exhaust side of the drive assembly. Importantly, the intake 92 and exhaust 94 tubes are of bellowed construction allowing the tubing to flex in response to vibrations of the drive assembly, without transferring the vibration to the housing base 24. The intake 92 and exhaust 94 tubes simply snap into the base plate 82 and are clamped (via clamps 98) to an associated nipple in the drive assembly.

Referring now to FIGS. 2, 4, 5, 12 and 14, the drive assembly generally includes a valve head 100, a cylinder 102, a piston 104 and an electromagnet motor 106, all aligned concentrically about a piston axis 108 (see FIG. 12). The entire drive assembly is bolted (via bolts 109) to a motor mount 110 and mounted to the base plate 82 via resilient mounts 112. The four resilient mounts 112 have narrowed necks between enlarged heads and bodies that slideably snap into four open ended slots 114 in the motor mount 110. Slightly enlarged bottom ends of the resilient mounts 112 can be pushed straight down through openings 116 in the base plate 82, thereby, the motor mount 110 is captured by the resilient mounts 112 and flexibly mounted to the base plate 82. The resilient mounts 112 are preferably made of a suitable rubber, such as an EPDM with a durometer of about 60, so as to provide enough stiffness to securely mount the drive assembly while allowing enough flexibility to isolate the vibrations of the drive assembly.

With primary reference to FIGS. 2, 4 and 12, the drive assembly will now be described in detail. The valve head 100 defines an intake side 120 and an exhaust side 122 each having a respective nipple 124 and 126 to which the intake 92 and exhaust 94 tubes connect via clamps 98. A valve plate 128 mates with the valve head 100. The valve plate 128 has a groove for a double D-shaped (bisected circle) o-ring 130 sealing against the valve head 100 to isolate the intake side 120 from the exhaust side 122. The valve plate 128 is generally disk-shaped and defines a pair of intake ports 131 (shown in phantom in FIG. 4) and a pair of exhaust ports 132. Each pair of ports is covered by respective thin metal flapper valves 134 and 135 (flapper valves 134 and 135 can be supported by valve stops (not shown)). The intake 131 and exhaust 132 ports are in communication with the associated sides of the valve head 100 and the inside of the cylinder 102, which fits into another groove in the back side of the valve plate 128 sealed with another o-ring 136 (see FIG. 13). The opposite side of the cylinder 102 fits around a hub of a retainer collar 138. The retainer collar 138 has four threaded openings, which receive four bolts 140 fit through four ears of each of the valve head 100 and valve plate 128 to clamp the cylinder 102 tightly together with these components.

The other side of the retainer collar 138 clamps one or more leaf springs 142 with a recessed groove (having alignment features as discussed below) in a spacer ring 144. The opposite side of the spacer ring 144 receives a stator 146 of the electromagnet motor 106. The stator 146 is a slotted annular member having a circular base and concentric inner 148 and outer 150 cylindrical walls (with axial slots 169 in outer wall 150), which define an annular channel 152 therebetween. A wire coil 154 is disposed in a bobbin 156 within the channel 152. The bobbin 156 has three posts that extend through openings in the base of the stator 146 and are engaged by retainers 153 to retain the bobbin 156 and coil 154. A diode (not shown) may be electrically coupled to the coil 154 to rectify the alternating current input signal so that it drives an armature (or shuttle) 158 in only one direction, preferably toward the stator 146. Conductive tabs 160 for coupling the coil 154 to the power are also included.

The armature 158 has a series of axial bores therethrough and slides in and out of a side (right in the drawing) of the stator 146 when the coil 154 is energized. The armature 158 has a short hub with an axial bore 162 that receives a bottom end of a connecting rod 164. The connecting rod 164 is suspended along the piston axis 108 by the leaf springs 142 and passes through the center bore in the stator 146. The connecting rod 164 is secured to the armature 158 and the piston 104 by a long bolt 166 threaded into the piston 104 and mounting a mass disk 168 under its head.

The stator 146 is clamped between the spacer ring 144 and another spacer ring 172. That spacer ring 172 clamps one or more additional leaf springs 142 against a second retainer collar 174. Four tie rods 173 extending through ears in the first retainer collar 138 are threaded into openings in ears of the second retainer collar 174 to unite the components of the motor 106. The retainer collars 138 and 174 also have threaded openings receiving bolts 109 to connect the motor mount 110 and thereby mount the entire drive assembly to the base plate 82 via the resilient mounts 112, as described above.

As shown in FIGS. 16 and 17, at the side opposite the cylinder 102, the retainer collar 174 (as well as spacer ring 144) has a circular groove 176 about its inner periphery with three alignment pockets 178, the sides of which taper asymmetrically away from the groove 176. One or more leaf springs 142 fit into the groove 176 and their asymmetrically tapered alignment tabs 180 fit into the pockets 178. Due to the asymmetric configuration of the pockets 178 and tabs 180, the leaf springs 142 can seat properly into the retainer collar 174 (and the spacer ring 144) in only one angular and axial orientation (relative to the piston axis 108). The alignment tabs 180 are spaced apart about 120 degrees so that the springs can be mounted in one of three angular orientations. If a single angular orientation is desired, the alignment tabs 180 could be spaced asymmetrically. This facilitates and ensures the assembly of multiple leaf springs 142 in the same orientation at both ends of the connecting rod 164. In particular, the web pattern (and arcuate slots) of the leaf springs 142 are aligned, and any curvature or bowing of the leaf springs 142 out of the plane perpendicular to the piston axis 108 (which can occur from the die cutting process) will be in the same direction for each leaf spring 142.

This is important to ensure that the motor 106 has the spring rate for which it was designed. Specifically, during development the pump is tuned to operate at a frequency at or near its natural resonant frequency. In particular, the pump is operated with a load applied and using a calculated spring-mass system (i.e., the combination of spring rate of the springs 142 and mass of the moving components, namely the piston 104, armature 158, connecting rod 164 and any mass disk 168). The frequency of the input signal to the motor 106 is varied as various parameters are measured. For example, because power consumption goes up as the input frequency strays from the resonant frequency of the spring-mass system, power consumption measurements can be used to adjust the spring-mass system so that its natural frequency will be at or near that of a typical input signal, for example 60 Hz. Operating the pump at the resonant frequency improves efficiency, and reduces vibration, and thereby noise. The spring-mass system can also be adjusted to operate efficiently at different pressures. For example, by increasing mass or spring rate the spring-mass system can be made to operate at or near resonant frequency while the pump is providing increased pressure output. It should be noted that the mass disk 168 is used as a cost effective alternative to increasing or decreasing the mass of the piston, armature and/or connecting rod.

As is well understood, the piston 104 is driven by movement of the armature 158, when energy is supplied to the wire coil 154, to reciprocate within the cylinder 102. The piston 104 has an enlarged head with a peripheral groove holding a split piston ring 170 that seals against the cylinder 102 when pressure is developed. The stroke length is approximately 8 mm (4 mm in each direction) and is positioned approximately 1 mm from the top of the cylinder when at top dead center.

Given the single cylinder arrangement of the pump 20, the reciprocating piston 104, armature 158 and connecting rod 164 can cause the drive assembly inside the housing to vibrate. The leaf springs 142 absorb much of the energy from these moving components. The number, size and thickness of the leaf springs 142 are selected to achieve a spring rate determined primarily according to the mass of the piston 104 and the input frequency. The leaf springs 142 are selected so that in combination (between the two stacks) they result in a resonant frequency of the piston 104 and springs 142 (i.e., the spring-mass system) approximately equal to the input frequency, which is typically 50 or 60 Hertz. For example, in one preferred embodiment there is a stack of two springs in the second retainer collar 174 and a stack of two springs in the spacer ring 144 near the piston 104. If the stroke length were to be increased, for example if the pump to be used in an application requiring more air flow, the springs 142 could be of a thinner gauge, in which case the number of springs may be increased to three in each stack to achieve the same spring rate.

With reference to FIGS. 5-7, 9-11 and 13, air flow through the pump will now be described in detail for a preferred compressor embodiment of the pump. When the drive assembly is operating, ambient air is drawn into the pump intake through the intake air passage 36 around the periphery of the top cover 28. As shown in FIGS. 5-7, the intake air is drawn through the intake air passage 36 flowing upwardly along the ribs 66 and makes its way into the outer cavity 50 through the cut-out 64 in the peripheral wall 62 of the filter tray 46. Intake air then moves from that end of the outer cavity 50 through the filter 48 and around the risers 60 in the filter tray 46 to the openings 58 in the filter tray 46 and the partition 44 where it enters the inner cavity 42. The pressure of the intake air drops as it passes through the small openings into the inner cavity 42. The inner cavity 42 is effectively larger than the outer cavity 50, which allows the intake air to expand. The expansion of the intake air in the inner cavity 42 helps to dissipate the pulsations in the intake air arising from the operation of the intake valve. After entering the inner cavity 42, the intake air turns through a bend of about 90-180 degrees and travels to the other end of the inner cavity 42 to the orifices 52 where it travels down the vent tubes 72 and into the interior of the housing shroud 26. As shown in FIG. 5, the vent tubes 72 are located to direct the intake air into the slots 169 in the stator 146 of the motor 106 so as to convectively cool the coil 154. After passing through and around the motor 106, the intake air passes through the openings 80 in the base plate 82 into the wide part of the intake chamber 86. The air is routed through the narrowed part and up through the intake opening 90 in the base plate 82 and the intake tube 92 to the intake side 120 of the valve head 100, as shown in FIG. 10. Reciprocation of the piston 104 draws air into the cylinder 102 and compresses it. The pressurized air is then passed through the exhaust side 122 of the valve head 100 and down through the exhaust tube 94 in the base plate 82 to the wide part of the exhaust chamber 88 through exhaust opening 96. As shown in FIGS. 9 and 11, the pressurized air flows through the narrow portion of the exhaust chamber 88 and out through the outlet opening where fitting 34 is attached to suitable hose or tubing (not shown).

Since the pump has only a single cylinder, the pressurized air is pulsed at the rate of the input frequency, for example 60 Hz. The inventors have determined that pulsations in the output air can adversely effect the operation of the pump. In particular, the output lines act as resonant chambers, having their own natural frequency. If the pulsations in the output air are at a different frequency, the air will effectively encounter increased resistance going through the output lines. This creates excessive back pressure on the pump so that the spring-mass system can be made to operate at a different (non-resonant) frequency, thereby decreasing the efficiency of the pump. This makes the pump more sensitive to variations in input frequency which can further decrease efficiency. By reducing the amplitude of the pulsations in the air before leaving the pump, this problem can be avoided.

To that end, as shown in FIGS. 3, 9 and 13, the pulsations in the pressurized output air leaving the pump are dampened by a diaphragm 190 in the exhaust chamber 88. In particular, the diaphragm 190 is preferably a rubber disk mounted to the top of a cup 192 defined by the housing base 24 in the wide part of the exhaust chamber 88 by a support ring 194 bolted (via bolts 198) to the base 24. The cup 192 defines a trapped air pocket 196 below the diaphragm 190. As the pressurized air flows through the exhaust chamber 88, the pulsations act against the diaphragm 190, tending to make it flex into the cup 192. The air pocket 196 will compress slightly, but tend to resist inward movement of the diaphragm 190. This reactive force on the diaphragm 190 will tend to counter, and thus cancel or reduce the amplitude of, the pulsations in the exhaust chamber air. The diaphragm 190 and trapped air pocket 196 act similar to an accumulator, and as a result, allow the pump to output smoother, relatively non-pulsed, air through the output lines. As a result pump inefficiencies are avoided that may otherwise arise from changes in the length or diameter of the output lines or from changes to the input frequency to the motor. As an example of one advantage, the inventors have determined that use of the diaphragm in this way allows a single mass-spring system to be used for both 50 and 60 Hz applications.

An illustrative embodiment of the present invention has been described above in detail. However, the invention should not be limited to the described embodiment. To ascertain the full scope of the invention, the following claims should be referenced. 

1. A linear pump comprising a cylinder and piston disposed along a piston axis and an electromagnet having a stator containing a wire coil driving an armature connected to the piston to reciprocate the piston within the cylinder along the piston axis, wherein the cylinder is in communication with an exhaust chamber through an exhaust port of a valve head, and wherein the exhaust chamber includes an attenuator movable in response to pressure fluctuations in the exhaust chamber to attenuate pulsations in an air stream exiting the exhaust chamber.
 2. The pump of claim 1, wherein the attenuator is a diaphragm.
 3. The pump of claim 2, wherein the diaphragm is mounted over a cavity allowing the diaphragm to move toward the interior of the cavity in response to pressure fluctuation in the exhaust chamber.
 4. The pump of claim 3, wherein the cavity is a circular recess.
 5. The pump of claim 2, further including a diaphragm support.
 6. The pump of claim 5, wherein the diaphragm support has a center opening.
 7. The pump of claim 1, wherein the diaphragm is a rubber disc.
 8. The pump of claim 1, wherein there is only one cylinder and piston arrangement. 